The present invention is directed to a method and apparatus for centrifugal testing pumps, and more particularly, a method and apparatus for determining pumping efficiency of large quantity centrifugal slurry pumps using air as a test medium to predict pump performance and efficiency. The use of air as a testing medium requires less power to operate the pump, and a open test loop consisting of a light construction flow tube, section pipe and discharge pipe sections making for a simpler, less costly, easier to use test rig.
Centrifugal slurry pumps are large quantity pumps, such as large dredge pumps, for moving a large volume of solid-liquid mixture. As described, for example, in Applicants commonly assigned U.S. Pat. No. 4,923,369, which issued May 8, 1990, a centrifugal-type pump consists basically of a rotatable impeller enclosed by a collector or shell. As the impeller is rotated, it generates velocity head at the periphery of the shell. The shell collects the velocity head and converts it to a pressure head. There are many configurations within the framework of this basic design. In one common configuration illustrated in FIG. 1, the flow enters the shell on one side along the axis of rotation of the impeller, that is, the flow enters the shell at a point adjacent to the center of the impeller, referred to as the xe2x80x9ceyexe2x80x9d of the impeller, while the discharge of the shell is located at a point tangent to the shell outer periphery.
The impeller is connected to a drive shaft 20, which protrudes away from the shell and is rotatably supported by suitable bearings blocks 21. A motor (not shown) rotates the shaft 20 and the impeller within shell. The usual packing (not shown) for surrounding shaft 20 in the central portion of the back side of the shell, prevents leakage.
The efficiency and performance of a centrifugal slurry pump is normally determined by pumping water around a test loop and recording the differential head across the pump, the flow through the pump and the pump input power. The head and power at constant rotational speed (rpm) varies with the flow as shown in FIG. 2. Head quantity performance at constant rpm has a characteristic curve that is usually established from ten or so sets of head measurements at different flows.
To determine performance of the pump, calculations for head and efficiency are made. The volume of liquid pumped is referred to as capacity and is generally measured in liters per second. The height to which liquid can be raised by a centrifugal pump is called total dynamic head (TDH) and is measured in meters. This does not depend on the nature of the liquid (its specific gravity) so long as the liquid viscosity is not higher than that of water. Water performance of centrifugal pumps is used as a standard of comparison because practically all commercial testing of pumps is done with water.
The head (H), or TDH as it is commonly called, is determined from the differential pressure across the pump and with appropriate velocity head and datum corrections using the well known Bernoulli""s Equation 1:                     H        =                  TDH          =                                                                      V                  B                  2                                -                                  V                  A                  2                                                            2                ⁢                g                                      +                                                            P                  B                                -                                  P                  A                                                                              ρ                  f                                ⁢                g                                      +                          (                                                z                  B                                -                                  z                  A                                            )                                                          (        1        )            
where
TDH is usually in meters of H2O
V is usually in meters/sec.
P is usually in Newton/meters squared (or Pascal)
xcfx81 is the density of water in kg/m3 
g is 9.81 meters/second squared
z is in meters of H2O.
The performance shown in FIG. 2 varies with speed according to well-established laws, which may be used to establish the hydraulic performance of the pump at other speeds. When testing a pump, it is normal to test at or near the expected pump operating speed. The following discussing will be limited to constant speed.
The power output of the pump is determined by the product of Q, which represents the quantity of fluid pumped (m3/sec) and H, and is given by:
(Power)out=xcfx81gQHxe2x80x83xe2x80x83(2)
This relation applies in any consistent system of units. Thus for SI units, equation (2) gives the power out in watts, which is usually divided by 1000 to obtain kilowatts. In the units in common use in the United States, Q is expressed in US gallons per minute, and H in feet and a numerical coefficient is required in the equation.
The input power to the pump can be defined as (T)(xcfx89), where T is the torque on the shaft of the pump and xcfx89 is the angular velocity. Using seconds as the unit of time, xcfx89 is equal to 2Πn where n is measured revolutions per second. In practice, the speed of rotation is usually measured in revolutions per minute, even where SI units are generally employed. The symbol N (or rpm) will be reserved for revolutions per minute, and thus:
xcfx89=2Πn=2ΠN/60xe2x80x83xe2x80x83(3)
and the input power is then defined as:                                           (            Power            )                    ⁢          in                =                              2            ⁢            π            ⁢                          xe2x80x83                        ⁢            nT                    =                      2            ⁢                          π              ⁡                              (                                  NT                  60                                )                                                                        (        4        )            
With torque in Newton-meters, this equation gives power in watts. In the United States, torque is generally measured in foot-pounds, and the input power is known, for historic reasons, as brake horsepower. When defined in this manner, a further numerical coefficient is required.
A final point, and in fact the main point concerning the pump""s input and output, is pump efficiency, denoted xcex7, which is the ratio of output power to input power, i.e.;                     η        =                                            (              Power              )                        out                                              (              Power              )                        in                                              (        5        )            
This relation applies in any system of units. For ideal pump efficiency, xcex7 is 1.00 or 100%. However, in practice, pumps necessarily have lower values. Efficiencies of over 90% can be achieved for large water pumps. Efficiencies of slurry pumps tend to be somewhat less.
The resulting measured efficiency of a normal slurry pump (at constant speed) has a characteristic (curve) as shown in FIG. 2. The highest value along the curve is usually called the Best Efficiency Point Efficiency (BEPE) and the flow (Q) and head (H) corresponding to the BEPE are referred to as the BEPQ and BEPH, respectively. The theoretical head of a centrifugal pump may be defined as:
Ht=[u2ct2xe2x88x92u1xe2x88x92ct1]/gxe2x80x83xe2x80x83(6)
where
u is the tangential velocity; and
ct the tangential component of the absolute flow velocity.
As losses have been disregarded, H is a theoretical head.
Equation (6) is often called the Euler equation, after its originator (Euler, 1756). The term u1ct1 refers to the flow entering the eye of the impeller. At the best efficiency point this term effectively reduces to zero. Thus it is ignored when considering the idealized machine with efficiency of 100%. The vector diagram at the exit of the impeller shows that:
c2t=u2xe2x88x92cm2 cot xcex2fxe2x80x83xe2x80x83(7)
where xcex2f is the angle between the relative velocity vector and the circumferential direction. It is somewhat smaller than the vane outlet angle. The term cm2 is the meridional component of outlet velocity (directed radially outward for most slurry pumps), which in turn is given by the discharge Q divided by the exit area of the impeller, i.e.:                               c          m2                =                  Q                      π            ⁢                          xe2x80x83                        ⁢                          D              2                        ⁢                          b              2                                                          (        8        )            
where b2 is the breadth between the shrouds at the outlet of the impeller.
All of the above are in terms of a fluid being pumped. Normally, this is thought of as some type of liquid as the fluid, or even air, noting then that the head produced value must also be in units of air. Centrifugal water and slurry pumps have been run on air. Provided the Reynolds (Re) number is high enough ( greater than 105), the head quantity performance data based on air as the flow medium has been found to be identical (within a normal commercial testing code tolerance) with that obtained on water. Recent tests carried out at the inventor""s test laboratory on a pump with an impeller diameter of 1.58 meters, for example, show that separate water and air tested head quantity values were within 1% of each other.
If a large slurry pump could have performance and efficiency tests carried out using air rather than a liquid, with sufficient accuracy, establishing pump performance could be done much easier and with less expense. The power necessary to perform the tests with air rather than a liquid such as water is reduced by the ratio of the density of water to air, i.e.; about {fraction (1/800)}th of the power when pumping on air rather than on water. This is because the fluid being pumped can be drawn from the surrounding atmosphere requiring less pipe work. Due to the significant reduction in generated pressures, the pump plate and piping components can be made lighter and less expensively.
However, performing tests with air requires extra precautions in order to obtain accurate head quantity measurements. The pressures used during such tests are also about {fraction (1/800)}th or so what would be normal as when using water. Thus, highly accurate pressure measurements must be taken, and the method of flow measurement requires a suitably constructed venturi meter with high physical accuracy calibrated for the corresponding Re-range.
While it is possible to test a pump using air and have this accurately predict the head quantity performance on water, the same is not necessarily true for obtaining an accurate prediction of pump efficiency. There are limitations associated with predicting pump efficiency using air. One problem associated with air testing is that while the different disk friction, impeller and leakage losses remain proportionally the same within the same physical components, the mechanical losses in the form of the bearing housing mechanical losses do not similarly scale with the change in pumping fluid. In other words, while all of the disk friction losses scale with the power out performance, the mechanical losses stay roughly the same. A plot of typical losses is shown in FIG. 3 which includes different centrifugal pump losses as a percent of input when pumping water. The specific speed, Ns, in FIG. 3, is a hydraulic number describing the design of a pump which is equal to:                               N          ⁢                      BEPQ                                    BEPH          0.75                                    (        9        )            
where BEPQ is the flow in GPM corresponding to the Best Efficiency Point Efficiency, and
BEPH is the head in feet corresponding to the Best Efficiency Point Efficiency.
It can be shown that the bearing housing losses when pumping water represent about 1% of the total power input to a centrifugal water pump. Tests, for example, on a 1.58 meter diameter impeller pump in the inventor""s test lab showed 1.03% of the power input was consumed by the bearing housing.
The resisting rotating torque of a normal bearing housing comprises the resistance of the bearings together with resistance of any seals. It can he shown that with most bearings, the resistance changes very little, with load being mostly dependent on the size of the bearings (and any seal), the temperature (and type) of the lubricant and the rotating speed.
The bearing houses for centrifugal slurry pumps are large, cumbersome units designed to support the heavy weight of the impeller and shell as well as the hydro-dynamical forces associated with pumping large quantities of slurry. When using air and the same bearing housing while running at the same speed, the hydraulic load is dramatically different than when using a liquid. However, the bearing housing losses remain roughly the same provided the temperature, lubricant type, speed, etc. are the same and assuming the bearings are designed to carry the radial and thrust loads directly.
In theory it would be possible to make an impeller that weighed about {fraction (1/800)}th of the original impeller and then design a special bearing housing for this weight and the air pumping loads. In practice, this is neither possible nor desirable so what remains is a bearing housing that is larger proportionally than that used in a water test.
On a 1000 kW input power pump, the normal pump bearing housing losses on water would be about 10 kW or 1%. The same pump with the same bearing housing, while pumping air, would require about                     1000        -        10            830        +    10    =      11.19    ⁢          xe2x80x83        ⁢    kW  
where 830 is the approximate relative density ratio of air to water and where the bearing housing losses are now,       10    11.19    =      89    ⁢    %  
of the total power.
Although it was previously mentioned that the bearing housing losses for air and water are roughly constant, they do, in fact, vary by about 1-1.5% depending on the temperature of the lubricant, etc. While this has a negligible effect on the performance of a pump being tested on water, its effect on the performance of a pump being tested an air is very significant. If, for example, the 10 kW value above becomes 15 kW, then the total power above on air would become 16.19 kW and the proportion of power absorbed by the bearing housing becomes 93%.
The mechanical losses in an air test are a disproportionately large portion of the losses on a pump and, as a consequence of that, with any small variation in them, it is not possible to calculate the pump efficiency with the required accuracy. In theory, it would be possible to reduce the size of the bearing housing for an air test on the basis of the reduction in input power. Unfortunately, as noted earlier, it is not normally possible or practical to reduce the weight of the impeller in the same ratio. As such, it is not possible or practical to manufacture a bearing housing for an air test that has a normal portion of losses. Different bearing housing temperatures result in variations in bearing housing torque which, on water, are negligible, but in an air test would at least cause a problem with test repeatability.
The present invention is directed to solving the problems described above with respect to air testing slurry pumps to predict performance and efficiency. According to the present invention, a test method and apparatus is disclosed to test a slurry pump using air as a test medium. The use of air as a test medium according to the present invention substantially reduces the costs associated with testing the pump, while ensuring an accurate prediction of pump efficiency in its intended operating environment.
Accordingly, one aspect of the present invention is directed to a method for determining pump efficiency using air as a test medium in place of a liquid. According to the method, an impeller test is performed wherein air is introduced into a pump having an impeller and a drive shaft axially mounted within a shell. The impeller test is performed at a substantially constant speed over a range of flows from zero to approximately 130% of the expected BEPQ flow for the pump to obtain the total torque values represented by the total power, or Ptot. Thereafter, the impeller is removed and the bearing house power losses are calculated using a dummy weight to simulate the weight of the impeller at the same speed used during the impeller test to obtain the total power loss associated with the bearing housing, or Pbrg. The power loss associated with the bearing housing, Pbrg, is then subtracted from the total power, Ptot, and the difference is multiplied by 1.01, which represents adding back into the equation the normal bearing housing power and stuffing box losses associated with water. The result is a corrected total input shaft power, or Pcor. Pump efficiency, xcex7, is then calculated based on the total power output (Pout) divided by the corrected total power input (Pcor). The pump efficiency figure obtained by this method provides an accurate prediction of pump efficiency within normal pump testing code.
According to another aspect of the present invention, an apparatus is disclosed for measuring pump efficiency using air as a test medium in place of water to save on costs associated with testing the pump. The apparatus includes a bearing housing for supporting an impeller drive shaft for a pump. The bearing housing is designed to support the normal weight of the structural parts of the pump but only a small fraction of the normal hydrodynamic force associated with the pump in normal operating conditions when pumping a fluid because air is being used as the pumping medium rather than a fluid. The apparatus further includes a torque measuring device and a variable speed motor. The bearings are small ball or roller bearings designed to support a smaller load than normal bearings used in the pump during normal operation. The use of smaller, constant friction bearings reduces the amount of power absorbed by the bearing housing. Because temperature of the bearing housing effects the torque required to rotate the drive shaft, a heating device is also included to heat and maintain the bearing housing at a relatively constant temperature.
These together with other objects of the invention, along with the various features of novelty which characterize the invention, are pointed out with particularity in the claims annexed to and forming a part of this disclosure. For a better understanding of the invention, its operating advantages and the specific objects attained by its uses, reference should be had to the accompanying drawings and descriptive matter in which there is illustrated preferred embodiments of the invention.
Other objects, advantages and novel features of the present invention will become apparent from the following detailed description of the invention when considered in conjunction with the accompanying drawings.